For reasons of comfort and performance, automated transmissions capable of power-shifting are increasingly used in utility vehicles. In this regard dual-clutch transmissions which shift without traction force interruption are particularly attractive. Transmission structures of dual-clutch transmissions for utility vehicles are known, for example, from DE 10 2005 033 027 A1 and DE 10 2006 054 281 A1. Such automated variable-speed transmissions comprise an intermediate transmission or a plurality of intermediate transmission groups arranged drive-technologically one after another, and if necessary a planetary transmission arranged drive-technologically downstream therefrom.
In a classical dual-clutch transmission, the gears are as a rule divided between two transmission branches in a direct and an indirect gear group, such that each group is associated with one clutch of the dual clutch. By means of the dual clutch a sequential shift sequence virtually without traction force interruption can be carried out, such that in each case the next gear is pre-selected in the currently load-free transmission branch and the gearshift takes place by overlapped actuation of the two clutches. The force flow of the gears can run in a conventional manner by way of a drive input shaft and a drive output shaft, or it can meander through the transmission by way of shafts that change a number of times.
Compared with pure power-shifting automatic transmissions of planetary design, automated variable-speed transmissions have efficiency and cost advantages. However, as the number of gears increases, so too do the structural size, the structural complexity and therefore the production cost. Since, depending on their field of use, utility vehicle transmissions require a relatively large number of gears in order to produce a defined transmission spread and to operate efficiently, particularly for use in utility vehicles it is expedient also to consider less expensive and more compactly designed, mixed transmission forms, namely so-termed partial dual-clutch transmissions. In such partial dual-clutch transmissions, besides a power-shifting transmission or transmission section with a dual clutch, a conventional transmission section that shifts with traction force interruption, for example a main transmission group or a downstream transmission that shifts with traction force interruption, for example a transfer box or an axle transmission, is also provided. With such partial dual-clutch transmissions, therefore, traction force interruptions have to be accepted in the case of some gearshifts.
For example, DE 10 2008 008 496 A1 shows a variable-speed transmission whose mode of action corresponds to that of a partial dual-clutch transmission. The variable-speed transmission has a first and a second input shaft which, in each case by way of one clutch of a dual clutch, can be connected to a drive engine. In addition there is a countershaft to which the two input shafts are coupled by a first and a second input gearset with different gear ratios. A drive output shaft can be coupled selectively to the countershaft by way of various further gearsets. A two-directional gearshift between a first gear and a second gear can be carried out without traction force interruption and without further shift processes by overlapped opening or closing of the two clutches, since the force flow in these gears only changes between the first and second input gearsets, but extends via the further gearset to the drive output shaft. The other gearshifts require shifting of the gear clutches involved, which results in a traction force interruption. To reduce the disadvantages of such traction force interruptions is the objective of another invention by the present applicant.
Another problem with partial dual-clutch transmissions and other automated transmissions having a dual clutch, is adapting the characteristic curve of the clutches, which are as a rule designed as friction elements. In this context the relationship between a clutch position or a regulating path point of a clutch and the coupling torque it can transmit is, as is known, an essential parameter, which is stored in the form of characteristic curves or performance graphs in a control computer and made available for the purposes of a transmission control process.
For example, if the clutch is open too far, starting processes can be delayed and the vehicle may even roll backward. On the other hand, if the clutch is closed too far, crawling may take place too actively and with unforeseen movement of the vehicle. During driving operation, if the characteristic clutch points (traction point, entrainment point, contact point, etc.) are not known accurately, shift processes may be accompanied by torque impulses and/or increased wear. The path and torque characteristics of a clutch can change due to temperature, rotational speed, wear and aging effects. Accordingly, for consistent, proper and comfortable control and operation of the dual clutch and the transmission, the relationship between the clutch position and its coupling torque must be updated regularly in a characteristic curve or table.
In accordance with known methods this relationship is determined by a teach-in process which essentially makes use of the parity between the engine torque and the clutch torque during slipping phases of the clutch. The adaptation can always take place when the clutch concerned is closed under load with slip. Basically, this is possible during starting processes or after gearshifts. Since in the case of dual-clutch transmissions the clutches are used in an overlapping manner during gearshifts, so that clear determination of the clutch torque is not possible, slipping phases of the clutch after gearshifts are useless for characteristic curve adaptation. Thus, fewer adaptation opportunities are available, in particular only the starting processes. Moreover, in starting processes in most cases the same gear and its associated clutch are used, so for the other clutch even regular starting processes are largely unavailable as adaptation opportunities. All in all therefore, this creates the problem that the association between clutch position and clutch torque cannot be determined for both of the clutches of a dual clutch with sufficient frequency and quality.